1. Field of the Invention
The present invention relates to a turbine nozzle suitable for reducing interblade loss caused between the nozzles and the moving blades of a steam turbine to improve the internal efficiency of the steam turbine.
2. Description of the Related Art
Techniques desirable for the performance improvement of steam turbines have been developed and applied successfully to practical steam turbines to achieve high operating efficiency. Noteworthy techniques effective in performance improvement are those relating to the improvement of internal efficiency. Those techniques are capable of being effectively applied to turbines of all kinds of turbine cycles operating under any fluid conditions and are noteworthy because of their capability of application to a wide range of application. Secondary flow loss among internal losses caused within a turbine is common to the many stages of an axial-flow turbine. The internal efficiency of a turbine is greatly dependent on measures taken to reduce secondary flow loss.
Incidentally, close examination of the shape and arrangement of blades is essential to the reduction of secondary flow loss attributable to vortices in the secondary flow generated in the nozzles. Recently developed advanced computer techniques capable of accurate analysis of three-dimensional flows have made possible the close examination and detailed three-dimensional analysis of the shape and arrangement of blades.
As an example of applications of the above-mentioned techniques, it has been known a turbine nozzle which comprises nozzle blades each of which is curved relative to the radial line passing the center axis of rotation of a steam turbine in a curve convex toward the fluid flowing direction with respect to a circumferential direction.
FIG. 6 is a fragmentary view of a part of a stage of an axial-flow turbine employing nozzle blades curved in the foregoing manner. Nozzle blades 1 are held between a diaphragm outer ring 2 and a diaphragm inner ring 3. In this nozzle, as a result of curvature of the blades in the above-mentioned manner, a velocity vector of the fluid flowing through a passage between the nozzle blades 1 is directed toward the diaphragm inner ring 3 in a root side area of the passage, and a velocity vector of the fluid flowing through a passage between the nozzle blades 1 is directed toward the diaphragm outer ring 2 in the tip side area of the passage. Such an action of the nozzle blades 1 suppresses the development of boundary layers on both sides walls of the diaphragm inner ring 3 and the diaphragm outer ring 2.
As another example of applications of the above-mentioned techniques, two types of turbine nozzles have been known which comprise nozzle blades which are arranged so that the ratio S/T, where S is the throat width which is the shortest distance between the trailing edge of the nozzle blade 1 and the back surface of another nozzle blade 1 adjacent to the former, and T is the pitch of the nozzle blades (see FIG. 7), is varied along the direction of the blade length to control flow distribution on the blade length for the improvement of the cascade performance.
One of said two types of turbine nozzles is shown in FIG. 8. Nozzle blades 1 shown in FIG. 8 are shaped and arranged so that the respective throat widths S1 and S3 at the root portion and the tip portion of the cascade are greater than the throat width S2 at the middle portion of the cascade to reduce the secondary flow loss in the vicinity of the side wall surface of the diaphragm inner and outer rings by increasing flow rates in the tip portion and the root portion of the fluid passage between the nozzle blades 1. This type of nozzle will be called a nozzle of a "three-dimensional design 1" hereinafter.
Another of said two types of turbine nozzles is shown in FIG. 9. Nozzle blades 1 shown in FIG. 9 are shaped and arranged so that the throat width S2 at the middle portions of the cascade is greater than the respective throat widths S1 at the root portion thereof and S3 at the tip portion thereof to reduce the secondary flow loss by increasing flow rates, relatively to in the root and tip portion of the fluid passage between the blades, in the middle portion thereof. In the middle portion of the fluid passage of the nozzle designed in above-mentioned manner, the flow of the fluid is not affected by the side wall surface of the diaphragm rings, hence secondary flow loss can be reduced. This type of nozzle will be called a nozzle of a "three-dimensional design 2" hereinafter.
As mentioned above, the performance of the cascade of the nozzle blades can be improved by three-dimensionally controlling the flow of steam by the nozzle blades disposed so that the ratio S/T varies along the length of the nozzle blades.
Interblade loss caused between the nozzle blades and the moving blades of the rotor of a steam turbine is one of the factors dominating the internal efficiency of the seam turbine. The interblade loss, in general, is the sum of unsteady loss and mixing loss, which will be described below.
Referring to FIG. 10 unsteady loss is caused by the passage of the moving blades (not shown in FIG. 10) through wakes. In other words, the unsteady loss is caused by the periodic variation of the inflow angle of the fluid relative to the moving blades due to the variation of velocity component of the fluid outflowing from the nozzle. The depth of the wakes decreases with the distance from outlet of the nozzle as measured in the direction of the flow, and unsteady loss decreases as the depth of the wakes decreases.
Mixing loss is caused by interference between streams of the fluid spouted into a free space. Mixing loss increases with the distance from the outlet of the nozzle in the flowing direction of the fluid. Accordingly, as shown in FIG. 11, the interblade loss .xi.3, i.e., the sum of the unsteady loss .xi.1 and the mixing loss .xi.2, reaches a minimum at a distance where a curve representing the former loss decreasing with distance L (see FIG. 12) in the direction of flow and a curve representing the latter loss increasing with the distance L in the direction of flow intersect each other. The distance where the interblade loss .xi.3 reaches a minimum is an optimum value of the distance L in the direction of flow.
Referring to FIG. 12, an optimum axial distance .delta.a between a nozzle blade 1 and a moving blade 4 is expressed by: EQU .epsilon.a=L.sub.opt sin .alpha..sub.2
where L.sub.opt is an optimum distance in the direction of flow, and .alpha..sub.2 is outlet flow angle of the nozzle. The "distance L in the direction of flow" is the distance between Line L.sub.n connecting trailing edges of the adjacent nozzle blades and Line L.sub.o connecting leading edges of the adjacent moving blades, as measured along the line inclined at inclination .alpha..sub.2 with respect to the Line L.sub.n. The axial distance .delta. is the distance between Line L.sub.n connecting trailing edges of the adjacent nozzle blades and line Lm connecting leading edges of the adjacent moving blades as measured in the axial direction.
In the aforementioned conventional nozzle, the outlet flow angle of the nozzle .alpha..sub.2 varies with longitudinal distance from root side end of the blade, as shown in FIG. 14, as a result of S/T variation along the length of the nozzle blades 1 as shown in FIGS. 8 and 9. The outlet flow angle of the nozzle also varies with distance along the length of the blade as a result of curvature of the nozzle blades 1 (FIG. 6).
Accordingly, the optimum axial distance .delta.a varies with the value of sin .alpha..sub.2 which varies with distance along the length of the blade as shown in FIG. 15. Hence, the internal efficiency of the turbine cannot be satisfactorily improved without optimization of the distance L in the direction of flow between the nozzle blades and the moving blades, even if the nozzle blades are curved circumferentially as shown in FIG. 6, or the S/T varies with distance along the length of the blade as shown in FIGS. 8 and 9, for reducing the secondary flow loss.